There are known highly efficient heat transfer devices or heat pipes featuring a conglomeration of very useful characteristics such as a rather low thermal resistance enabling to transfer high density heat fluxes at a small temperature differential between a heat source and a heat sink, low weight per unit heat transferred, high reliability due to absence of moving parts, moderate overall dimensions and the capability of being employed in a wide range of temperatures. In addition, heat pipes can be made in a broad variety of shapes and sizes for special heat transfer situations.
Structurally, the conventional heat pipe is very simple. It is normally a pressure-tight vessel fabricated as a rule from metal, the atmospheric air being removed from the interior of the vessel. The inner surface of such a vessel is lined with a capillary material wet by a liquid which functions as a heat transfer fluid.
Operation of the heat pipe is based on the well known laws of physics. When heat is applied from a heat source to one end of the heat pipe, the heat transfer fluid is caused to evaporate from the capillary material to absorb the latent heat of vaporization, whereby vapor is moved toward the other (cooled) end of the heat pipe to condense therein for the heat of condensation to be transferred to an outer heat sink through heat conduction. The thus condensed heat transfer fluid is absorbed by the capillary material to be moved back by virtue of a capillary pressure head toward the evaporation zone thereby completing the working cycle of the heat pipe. High efficiency of the heat pipe as a "heat conductor" is therefore determined by that liquids feature rather high heat of vaporization which enables to remove from the evaporation zone considerable heat fluxes at a relatively low consumption of the heat transfer fluid, as well as by that heat is transferred mainly by vapor which is moved along a pipe without the need for a high pressure differential since the hydraulic diameter of the vapor passages is, as a rule, sufficiently great.
A principle equation associated with heat pipe operation is based on the balance of pressures and may be expressed as: EQU .DELTA.Pc.gtoreq..DELTA.Pb+.DELTA.Pv+.DELTA.Pg (1),
where
.DELTA.Pc is the capillary pressure head, in N/m.sup.2 ; PA1 .DELTA.Pb is the pressure differential in the liquid moving in the capillary material, in N/m.sup.2 ; PA1 .DELTA.Pv is the pressure differential, of the vapor in the vapor passage, in N/m.sup.2 ; and PA1 .DELTA.Pg is the hydrostatic head determined by the mutual interposition of the evaporation and condensation zones of the heat pipe in a mass force field, in N/m.sup.2. PA1 .theta. is the value of the extreme angle at which the inner wall of the capillary pore is wet by the liquid, in deg. PA1 .eta.e is the dynamic viscosity factor, in N.multidot.s/m.sup.2 ; and PA1 .zeta. is the effective length of the heat pipe, in m; and PA1 .rho.e is density of the liquid, in kg/m.sup.3. PA1 .rho.l is the density of the heat transfer fluid in a liquid phase, in kg/m.sup.3 ; PA1 g is the acceleration of gravity, in m/s.sup.2 ; and PA1 .phi. is the angle between the longitudinal centerline of the heat pipe and the horizontal, in deg. PA1 P.sub.1 is the pressure of vapor above the evaporation surface of the vapor release passages, in N/m.sup.2 ; PA1 T.sub.1 is the temperature of vapor in the vapor release passages, in K; PA1 .DELTA.T is the vapor temperature difference between the evaporation surfaces, in K; and PA1 R is the universal gas constant, in J/K.multidot.kg.
In its simplest form the capillary pressure head for capillary pores of generally cylindrical form may be expressed by the Laplace equation: ##EQU1## where .sigma. is the surface tension coefficient, in N/m;
This equation is true if the curvature radius of the vapor-liquid interface in the vapor condensation zone tends to be infinite, which corresponds to a flat interface, or if the wetting angle in the condensation zone is 90.degree..
When capillary pores are of complex configuration the capillary pore radius is substituted by a notion of effective radius which can be found experimentally.
Pressure differential in a laminar flow of incompressible viscous liquid moving through a cylindrical capillary pore having a radius of r.sub.c may be described by the Hagen-Poiseuille formula: ##EQU2## where G is the mass consumption of liquid, in kg/cm;
The movement of vapor in the heat pipe is governed by more complex laws and may vary in the evaporation zone, condensation zone and the transport (adiabatic) pipe portion. Therefore, the complete pressure differential in the vapor phase .DELTA.Pv is generally the total of pressure differentials at the above three portions of the heat pipe. Because a detailed analysis of the component pressure losses in the vapor phase is outside the scope of the present invention, it is suffice to cite a publication entitled "Heat Pipes" by P. D. Dunn and D. A. Reay, Pergamon Press, Oxford, New York, Toronto, Sydney, Paris, Braunscheweig, 1976, where such an analysis is contained in pages 35 to 49.
The last term of the equation (1) determined by the hydrostatic head of the liquid is determined through: EQU .DELTA.Pg=.rho.l.multidot.g.multidot..zeta..multidot. Sin .phi.(4),
where
Depending on the mutual position of the evaporation and condensation zones in the field of mass forces, the term .DELTA.Pg of the equation (1) enters this equation either with a positive sign (+) or with a negative sign (-). When the evaporation zone is above the condensation zone, the angle of inclination of the heat pipe is considered positive, while Sin .phi.&gt;0 and .DELTA.Pg has a positive sign (+) imparting a hydrostatic resistance. In consequence, an increase in the length of the heat pipe and in the angle of inclination thereif result in an increased hydrostatic pressure attaining its maximum value at .phi.=90.degree.. The hydrostatic pressure .DELTA.Pg contributes to a great extent to the total of pressure losses. Therefore, it must be taken into account even at negligeable inclination angles of the heat pipe, as well as at horizontal positioning of heat pipes of larger diameter. Especially susceptible to variation in the positive inclination angle in the field of mass forces are low-temperature heat pipes wherein use is made of heat transfer fluids having a relatively low surface tension factor. In this case, it is advisable to employ capillary materials having small radius of capillary pores in order to attain sufficiently high values of the capillary pressure head .DELTA.Pc. However, according to the expression (3), the growth in the hydraulic resistance is directly proportional to the square of the pore radius. This results in that the distance over which heat is transferred and the amount of heat flux are limited to such an extent that their advisability is questionable when rated operating conditions include situations where the heat pipe may be oriented such that the liquid phase of the heat transfer fluid must move against a gravity head or other mass forces.
There is known a heat pipe construction described in U.S. Pat. No. 3,666,005. This heat pipe is made up of a plurality of interconnected serial sections, each of the sections being actually an independent heat pipe. The inner surface of the sections is lined with capillary material saturated with a heat transfer fluid. The sections are so interconnected that an end face wall confining the condensation zone of a serially preceding section is integral with an end face wall confining the evaporation zone of every succeeding section, and so forth.
Therefore, the arrangement is such that the condensation zone of every preceding section is in thermal contact with the evaporation zone of the succeeding section of this heat pipe assembly. Because the heat transfer fluid is calculated independently in each of the sections and the length of each such section is relatively small, it stands to reason that within each of the sections the distance over which the liquid heat transfer fluid has to travel through the capillary material is rather short, which makes it possible to use capillaries with sufficiently large radius to enable to transfer markedly larger heat fluxes with the heat transfer fluid travelling against a gravity head as compared to conventional heat pipes.
However, this known heat pipe has a high thermal resistance caused by that heat transfer between the sections is effected by virtue of heat conduction through the separating walls possessing a certain amount of resistance to heat. Apparently, in order to increase the overall length of such a heat pipe, it is necessary to employ larger number of sections. In consequence, this leads to a greater number of walls separating the sections the total heat resistance of which makes up the overall thermal resistance of the heat pipe. It can be easily assumed that the thermal resistance of a heat pipe made up of a plurality of such sections will be higher than the thermal resistance of conventional heat pipes whereby the basic advantage of a heat transfer apparatus of this type, such as low thermal resistance, will be lost. Therefore, at a given temperature difference between a heat source and a heat sink the heat flux capacity of the abovedescribed heat pipe will be lower than that of conventional heat pipes.
Attempts to increase the heat flow transferred through reducing its hydraulic resistance resulted in a heat transfer apparatus protected by U.S. Pat. No. 3,741,289. This heat transfer apparatus is fashioned as a closed conduit defining an essentially circular heat link comprising at one portion thereof a vaporizer of capillary material saturated with a heat transfer fluid in thermal contact with a source of heat. A portion of the conduit remote from the vaporizer is adapted to maintain thermal contact with the heat sink. A portion of the conduit adjacent the vaporizer is provided with a liquid header. One part of the conduit disposed between the heat source and the heat sink serves to transmit the heat transfer fluid in a vapor phase, while the other part thereof is intended to carry the heat transfer fluid in a liquid phase. The apparatus is capable of providing a contact of the heat transfer fluid in a liquid phase with the vaporizer under no heat load. To this end, there is provided a reservoir arranged away from the heat link and communicating with the apparatus by way of a passage. The reservoir has a flexible diaphragm separating the heat transfer fluid from another heat transfer fluid partially in a liquid and partially in a vapor state the vapor pressure of which fluid exerted on the vaporizer under zero heat load is higher than the vapor pressure of the first heat transfer fluid and, conversely, it is lower when the temperature of vapor of the heat transfer fluid is raised subsequent to the application of a heat load. Therefore, in the absence of heat load the diaphragm assumes a curved or arched position toward one side of the reservoir for the heat transfer fluid to be driven from the reservoir to come into thermal contact with the vaporizer. When the pressure and temperature of vapor released by the heat transfer fluid subsequent to the application of the heat load have been increased, the heat transfer fluid is driven from the vapor portion of the conduit into the liquid portion thereof to come into contact with the outer surface of the vaporizer through the liquid header. Excess heat transfer fluid is forced into the reservoir to cause the diaphragm to assume a position curved toward the other side of the reservoir.
High heat flux capability of this apparatus is assured by that the distance travelled by the heat transfer fluid in the capillary material toward the evaporation surface is relatively small. Therefore, pressure losses in this apparatus are much less than in conventional heat pipes, which in turn enables to reduce the effective radius of the capillary pores and thereby increase the capillary head providing a motive force for the heat transfer fluid.
However, inherent in the above heat transfer apparatus is, in the first place, a disadvantage residing in a relatively small surface area intended for carrying the heat transfer fluid toward the vaporizer occupying a narrow annular portion of its outer surface. Extending the length of the vaporizer surface to convey a heat load thereto may cause insufficient feeding of remote portions of the vaporizer due to capillary resistance and, as a result, to essentially the same limitations in the travel of the heat transfer fluid against the action or mass forces as in conventional heat pipes. A second disadvantage resides in the overall bulk of the apparatus due to the liquid header and the separate reservoir arranged outside the heat link. Thirdly, the apparatus may have insufficient reliability because the movable element thereof, i.e. the diaphragm, is susceptible to residual deformations and mechanical wear.
A further reduction in the hydraulic resistance at a portion of the travel path of the heat transfer fluid in a vapor phase through the capillary material has been attained in a construction of a heat transfer apparatus according to USSR Inventor's Certificate No. 691,672.
This known apparatus comprises evaporating and condenser chambers communicable through conduits, the first of the conduits being intended to convey the heat transfer fluid in a vapor phase, the second conduit serving to carry the heat transfer fluid in a liquid phase. Accommodated in the interior of the evaporating chamber coaxially therewith is a vaporizer of capillary material saturated with the heat transfer fluid and adapted to maintain a thermal contact with a heat source. The vaporizer consists of two portions end surfaces of which are tightly adjacent therebetween. Each portion of the vaporizer is provided with longitudinal and radial vapor release passages communicable with a vapor header incorporated into the vaporizer and having the form of an annular recess occupying a border area between the two portions of the vaporizer. The vaporizer further has a longitudinal axial passageway communicable with each of two end cavities defined by the end surfaces of the vaporizer and the walls of the evaporating chamber. Provided in the side wall of the evaporating chamber is an inlet port for a first pipe to communicate with the vapor header, whereas an end face wall facing the condenser chamber has an outlet port for a second pipe to communicate with the end cavity of the evaporation chamber, this outlet port of the second pipe being arranged either in said cavity or in the longitudinal axial passageway of the vaporizer.
The condenser chamber is generally a shell in the form of a cup the bottom of which faces the evaporation chamber. Installed inside the cup coaxilly therewith is another shell to form between the side and end surface of the first shell facing the evaporating chamber and respective surfaces of the second shell an annular space and a planoparallel space located at a right angle relative to the first space, the two spaces defining the interior of the condenser chamber. The end face wall of the first shell facing the evaporating chamber has an outlet port for the first pipe communicable with the interior of the condenser chamber, an inlet port for the second pipe being arranged in the side wall of the first shell to communicate with the interior of the condenser chamber and spaced from the first port lengthwise of the chamber.
The heat transfer apparatus is charged with a heat transfer fluid in the amount sufficient to saturate the vaporizer, fill the second pipe, a portion of the condenser chamber, the longitudinal axial passageway, one end cavity and partially the other end cavity.
During operation of the apparatus under heavy operating condition when it is oriented in the field of mass forces essentially vertically for the evaporation chamber thereof to overly the condenser chamber, substantial inconveniencies may arise due to occurence of a hydrostatic resistance .DELTA.Pg tending to reach its maximum value. In the absence of heat load the vaporizer is saturated with the heat transfer fluid, while the balance of the heat transfer fluid occupies a certain level in the pipes as in communicating vessels. When a heat load is applied to the vaporizer, the heat transfer fluid is caused to evaporate from the surface of the vapor release passages, the surface of the longitudinal axial passage and from the end surfaces of the vaporizer. However, thanks to the thermal resistance of the layer of capillary material saturated with the heat transfer fluid which separate said surfaces, a temperature difference and, consequently, a pressure difference occur in the region above these surfaces.
This pressure difference may be determined according to the Clausius-Clapeyron equation as follows: ##EQU3## where L is the latent heat of vaporization, in J/kg;
Under the action of this pressure difference the heat transfer fluid in a liquid phase is caused to be driven out from the first pipe of the condenser chamber to occupy the end cavities and the longitudinal axial passageway of the vaporizer wherefrom it moves essentially in a radial direction through the vaporizer to be conveyed in the evaporation surface of the vapor release passages.
In consequence, two levels of the liquid heat transfer fluid are established in the apparatus, particularly one in the upper end cavity at a temperature of T.sub.2 of vapour thereabove, and the other one in the condenser chamber at a temperature of T.sub.3 of vapor above this level. Therewith, it is necessary that a condition T.sub.3 &gt;T.sub.2 and P.sub.3 &gt;P.sub.2 be compiled with. This condition is fulfilled because a cooled heat transfer fluid is admitted to the evaporating chamber, the saturated vaporizer maintaining its function of a "thermal gate". It should be noted here that the temperature T.sub.3 is somewhat lower than the temperature T.sub.1 due to losses caused by the travel of vapor along the first pipe and the annular space of the condenser chamber, whereas the condition P.sub.3 &gt;P.sub.2 is realized in case the capillary pressure head in the vaporizer meets the following condition: EQU .DELTA.Pc.gtoreq.P.sub.3 -P.sub.2 +.DELTA.Pl+.alpha.Pv (6).
It is therefore evident that the pressure difference P.sub.3 -P.sub.2 is approximately equal in value to the hydrostatic pressure .DELTA.Pg which is exerted by a column of the liquid heat transfer fluid confined between free surfaces in the evaporation and condenser chambers.
Accordingly, since the distance travelled by the liquid heat transfer fluid in the capillary material is relatively short and not dependent on the length of both the heat transfer apparatus and the vaporizer per se due to the predominantly radial path of travel thereof, it becomes possible to employ capillary pores very small in radius. This affords to obtain a high capillary pressure head even when a heat transfer fluid with a relatively low surface tension factor is used. In addition, the aforedescribed apparatus is reliable and moderate in size because the end cavities function as a reservoir for the excess heat transfer fluid, while the moving parts are absent. The level of the heat transfer fluid is controlled by the heat transfer fluid itself through variations in the values of P.sub.2 and P.sub.3.
Among disadvantages inherent in the abovedescribed heat transfer apparatus are, firstly, the complicated arrangement of the system of vapor release passages which must be great in number to provide a sufficiently large evaporation surface, as well as the inconveniences in terms of providing a reliable and tight connection and accommodation of the two parts of the capillary vaporizer in the housing. Secondly, the insufficient evaporation surface of the vaporizer defined by the side surfaces of the radial vapor release passages which, as has been noted, cannot be numerous enough for purely technological considerations. This hampers vapor release to result in pressure losses therein. Thirdly, another disadvantage is the location of the outlet port of the second pipe in the evaporating chamber at below the level of the liquid heat transfer fluid in the upper end cavity when the apparatus is oriented at angle of inclination .phi.&gt;0.degree., which fails to allow the admission of the "cold" heat transfer fluid directly to the upper end cavity through the longitudinal axial passageway having a cross-section by far larger than the cross-section of the second pipe due to the deceleration in the travel velocity of the heat transfer fluid, as well as because of a direct thermal contact thereof with the walls of the longitudinal axial passageway, which result in an increase in the temperature of the heat transfer fluid. In consequence, the temperature T.sub.2 and the pressure P.sub.2 tend to grow in value leading to a corresponding increase in the values T.sub.1, T.sub.3 and P.sub.1, P.sub.3 and, accordingly, to increased temperature of the heat source wherefrom the heat transfer apparatus draws away heat. Fourthly, the provision of the narrow annular space in the condenser chamber having a hydraulic resistance tending to increase further due to a film of condensate flowing downwards and impaired convection when heat is transferred from the outer surface of the condenser chamber to the outside are also disadvantageous because they tend to reduce the highest heat flux capacity transferred by the abovedescribed apparatus.